U.S. patent number 4,196,157 [Application Number 05/922,454] was granted by the patent office on 1980-04-01 for evaporative counterflow heat exchange.
This patent grant is currently assigned to Baltimore Aircoil Company, Inc.. Invention is credited to Edward N. Schinner.
United States Patent |
4,196,157 |
Schinner |
April 1, 1980 |
Evaporative counterflow heat exchange
Abstract
A counterflow evaporative heat exchanger comprises a vertical
conduit in which a coil assembly is positioned. A fluid to be
cooled or condensed passes through the coil assembly while water is
sprayed downwardly over it and air is blown upwardly through it.
The coil assembly comprises a plurality of tubes which at each
level in the assembly are spaced apart from each other in the
horizontal direction by an amount greater than the diameter of the
tubes. This has been found to improve heat transfer in comparison
to counterflow evaporative heat exchangers using closely packed
coil assemblies.
Inventors: |
Schinner; Edward N. (Silver
Spring, MD) |
Assignee: |
Baltimore Aircoil Company, Inc.
(Jessup, MD)
|
Family
ID: |
25447070 |
Appl.
No.: |
05/922,454 |
Filed: |
July 6, 1978 |
Current U.S.
Class: |
261/155; 62/310;
261/DIG.11; 261/30; 62/305; 165/900; 261/29 |
Current CPC
Class: |
F28D
5/02 (20130101); Y10S 165/90 (20130101); Y10S
261/11 (20130101) |
Current International
Class: |
F28D
5/00 (20060101); F28D 5/02 (20060101); B01F
003/04 () |
Field of
Search: |
;261/30,36R,109-112,152-154,29,DIG.11,DIG.77 ;55/257R,259
;165/DIG.1 ;62/2,156,282,305,310 ;122/510 |
References Cited
[Referenced By]
U.S. Patent Documents
Other References
Imeco, Inc., "Evaporative Condensers", Bulletin S-160.1, Feb.,
1972..
|
Primary Examiner: Chiesa; Richard L.
Attorney, Agent or Firm: Sudol, Jr.; Michael C.
Claims
What is claimed and desired to be secured by Letters Patent is:
1. An evaporative counterflow heat exchanger comprising a conduit
of generally uniform cross section extending in a vertical
direction, a coil assembly positioned inside said conduit, said
coil assembly comprising inlet and outlet manifolds and a plurality
of tubes connected between the manifolds with different segments of
the tubes extending generally horizontally across the conduit in
equally spaced relation to each other at different levels in the
conduit, means maintaining said tubes spaced apart from each other
by an amount such that the spaces between adjacent tubes at each
level are each greater, by a finite amount, than the diameter of
said tubes but are less than twice the tube diameter, liquid
distribution means arranged in said conduit above said coil
assembly to distribute liquid down through said conduit and over
said coil assembly, fan means arranged to move a gas up through
said conduit between said tube segments in counterflow relationship
to said liquid at a velocity sufficient to entrain liquid from said
coil assembly and carry said liquid up past said liquid
distribution means, and mist eliminator means extending across
substantially the entire cross section of said conduit above said
liquid distribution means.
2. An evaporative counterflow heat exchanger according to claim 1
wherein the tubes in adjacent levels are staggered horizontally
with respect to each other.
3. An evaporative counterflow heat exchanger according to claim 1
wherein said levels are also separated by a distance at least as
great as the tube diameter.
4. An evaporative counterflow heat exchanger according to claim 1
wherein each tube extends back and forth across said conduit in a
serpentine manner in a vertical plane between a common upper
manifold and a common lower manifold.
5. An evaporative counterflow heat exchanger according to claim 4
wherein laterally adjacent tubes are staggered vertically with
respect to each other to produce horizontal staggering of the tube
segments at adjacent levels.
6. An evaporative counterflow heat exchanger according to claim 1
or 4 or 5 wherein said coil assembly includes vertically extending
spacer elements positioned between the adjacent tubes to space them
horizontally from each other.
7. An evaporative counterflow heat exchanger according to claim 6
wherein said spacer elements are squeezed between and frictionally
held in place by said tubes.
8. An evaporative counterflow heat exchanger according to claim 1
wherein said fan means is of a size capable of blowing gas up
through said conduit at a rate of at least four hundred feet (122
meters) per minute.
9. An evaporative counterflow heat exchanger according to claim 1
wherein said fan means is of a size capable of blowing air at a
velocity of about one thousand feet (305 meters) per minute in the
vicinity of said tubes.
Description
BACKGROUND OF THE INVENTION
1. Field of the Invention
This invention relates to evaporative heat exchange of the type in
which a fluid to be cooled or condensed passes through an array of
tubes while a liquid and a gas pass in counterflow relationship
over the outer surfaces of the tubes.
2. Description of the Prior Art
Counterflow evaporative heat exchangers are shown and described in
U.S. Pat. Nos. 3,132,190 and 3,265,372. Those heat exchangers
comprise an upwardly extending conduit containing an array of tubes
which form a coil assembly. A spray section is provided in the
conduit above the coil assembly to spray water down over the tubes;
and a fan is arranged to blow air into the conduit near the bottom
thereof and up between the tubes in counterflow relationship to the
downwardly flowing sprayed water. Heat from the fluid passing
through the coil assembly tubes is transferred through the tube
walls to the water sprayed down over the tubes; and the upwardly
flowing air causes partial evaporation of some of the water and
transfer of heat from the water to the air. The thus heated air
then flows upwardly and out from the system. The remaining water
collects at the bottom of the conduit and is pumped back up and out
through spray nozzles in recirculatory fashion.
The following additional United States patents also disclose
countercurrent or crosscurrent liquid-gas evaporator type heat
exchangers as thus broadly described: U.S. Pat. Nos. 712,704,
2,076,119, 2,228,484, 2,454,883, 2,680,599, 2,840,352, 2,933,904
and 3,996,314.
None of the foregoing patents is concerned with the particular
arrangement of tubing which makes up the coil assembly; and the
effect of the arrangement or positioning of the tubes on heat
transfer is not discussed in those patents. In most cases the tube
arrangements are shown only schematically or incompletely. In
general, however, the coil assembly tubes were packed into as tight
an array as possible to maximize the tube surface area available
for heat transfer. A tightly packed coil assembly also maximizes
the velocity of the air flowing between adjacent tube segments. The
resulting high relative velocity between the air and water promotes
evaporation and thereby enhances heat transfer. In several of the
above identified patents the surface area of the coil assembly
tubes is further increased by the use of closely spaced fins which
extend outwardly, in a horizontal direction, from the surface of
the tube segments.
There are other evaporative type heat exchangers in which the
liquid and gas flow in the same direction over the coil assembly.
Examples of these other devices, which are generally referred to as
co-current flow heat exchangers, are shown in U.S. Pat. Nos.
2,752,124, 2,890,864, 2,919,559, 3,148,516 and 3,800,553. In the
systems of the first four patents, the cooling coil tubes are shown
spaced apart but no indication is given that the coil spacing has
any effect on heat transfer. In any event it has been recognized in
the prior art that heat transfer in an evaporative heat transfer
device, whether of the co-current flow type or the countercurrent
flow type, is directly related to the total surface area of the
heat exchange tubes carried in the apparatus. Accordingly, it has
been the teaching of the prior art that where heat transfer was to
be maximized, the heat exchange tubes should be packed together as
tightly as possible.
In the system of the last mentioned patent the cooling tubes are
shown closely packed.
SUMMARY OF THE INVENTION
The present invention has for an object to increase the net amount
of heat transfer per unit area of cooling tube surface in a
counterflow evaporative heat exchanger.
The invention has for another object the lowering of construction
costs of a counterflow evaporative heat transfer device without any
corresponding reduction in heat transfer capability and without any
increase in operating costs.
It is a still further object of the invention to improve the
cleanability of cooling coil tubes in a counterflow evaporative
heat exchanger.
The present invention achieves these objects in a novel manner.
According to one aspect of the present invention, the number of
tubes in the coil assembly of a counterflow heat exchanger is
reduced from that which would previously have been considered
necessary to provide maximum heat transfer area and maximum gas
flow velocity. More specifically, according to this aspect of the
invention, the coil assembly in a counterflow type evaporative heat
exchanger is arranged in a conduit up through which a gas, such as
air, is blown and down through which a liquid, such as water, is
sprayed or otherwise distributed. The coil assembly is made up of
arrays of substantially equally spaced apart tube segments located
at different levels in the coil assembly region of the conduit. The
tube segments are spaced apart horizontally at each level by an
amount such that the space between adjacent tubes is greater than
the diameter of the tube segments but not substantially greater
than twice their diameter. Thus at each level in the conduit the
portion of the coil assembly occupied by tube segments is less than
fifty percent but not substantially less than twenty five
percent.
The above described spacing of tube segments results in a
significant reduction in the amount of tubing used in comparison to
prior art tightly packed coil assemblies; and accordingly the cost
of the coil assembly of the present invention is correspondingly
reduced from such prior art coil assemblies. Although this
reduction in coil assembly tubing is accompanied by a corresponding
reduction in tube heat transfer surface area, it has been found,
surprisingly, that the heat transfer for each unit area of the
cooling tubes is actually increased; and where the number of tube
segments is such as to occupy approximately forty percent of the
coil assembly cross section at each level in the heat exchanger,
the overall heat transfer capacity of the heat exchanger is also
increased.
According to a further aspect of the present invention, counterflow
evaporative heat transfer is carried out by spraying water down
over an assembly of tubes and blowing air up between the tubes
while a fluid to be condensed or cooled flows through the tubes.
The water is sprayed at a rate sufficient to form water films on
the tube. The air, in turn, is blown upwardly at a velocity in the
vicinity of the tubes sufficient to shear water from the films but
insufficient to strip the films completely from the tubes. More
specifically, the air velocity in the vicinity of the tubes is
maintained at more than four hundred feet (122 meters) per minute
but less than fourteen hundred feet (427 meters) per minute.
Preferably the air velocity is maintained at about one thousand
feet (305 meters) per minute in the vicinity of the tubes. Water
sheared from the films is entrained in the upwardly flowing air but
after the air leaves the tubes it passes through mist eliminators
which recover the water and redirects it back over the tubes.
There has thus been outlined rather broadly the more important
features of the invention in order that the detailed description
thereof that follows may be better understood, and in order that
the present contribution to the art may be better appreciated.
There are, of course, additional features of the invention that
will be described more fully hereinafter. Those skilled in the art
will appreciate that the conception on which this disclosure is
based may readily be utilized as the basis for the designing of
other arrangements for carrying out the several purposes of the
invention. It is important, therefore, that this disclosure be
regarded as including such equivalent arrangements as do not depart
from the spirit and scope of the invention.
BRIEF DESCRIPTION OF THE DRAWINGS
Selected embodiments of the invention have been chosen for purposes
of illustration and description, and are shown in the accompanying
drawings, forming a part of the specification, wherein:
FIG. 1 is a side elevational view, partially in section of a
counterflow evaporative type liquid-gas heat exchanger according to
the present invention;
FIG. 2 is a front elevational view, partially broken away and
partially in section, of the heat exchanger of FIG. 1;
FIG. 3 is a view taken along line 3--3 of FIG. 2, partially broken
away, and showing a coil assembly used in the heat exchanger;
FIG. 4 is a view taken along line 4--4 of FIG. 3, and partially
broken away;
FIG. 5 is a fragmentary perspective view showing a tube segment
array forming one portion of the coil assembly of FIGS. 3 and
4;
FIG. 6 is a diagrammatic repesentation of a view taken along line
6--6 of FIG. 5;
FIG. 7 is a view similar to FIG. 6 but showing a prior art tube
segment array;
FIG. 8 is a graph showing comparative heat transfer characteristics
of the present invention; and
FIG. 9 is a view similar to FIG. 1 but showing a modification of
the heat exchanger.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
The heat exchanger shown in FIGS. 1-6 comprises a generally
vertical conduit 10 of sheet metal construction and having, at
different levels in the interior thereof, an upper mist eliminator
assembly 12, a water spray assembly 14, a coil assembly 16, a fan
assembly 18 and a lower water trough 20.
The vertical conduit 10 is of rectangular, generally uniform,
cross-section and it comprises vertical front and rear walls 24 and
22 (FIG. 1) and vertical side walls 26 and 28 (FIG. 2). A diagonal
wall 30 extends downwardly from the front wall 24 to the bottom of
the rear wall 22 to define the water trough 20. The fan assembly 18
is positioned behind and below the diagonal wall 30. The fan
assembly comprises a pair of centrifugal fans 32 each of which has
an outlet cowl 34 which projects through the diagonal wall 30 and
into the conduit 10 above the water trough 20 and below the coil
assembly 16. As shown in FIG. 2, the fans 32 share a common drive
axle 36 and this axle is turned by means of a drive pulley 38
connected through a belt 40 to a drive motor 42.
A recirculation line 44 is arranged to extend through the side wall
26 of the conduit 10 near the bottom of the trough 20. The
recirculation line extends from the trough 20 to a recirculation
pump 46 and from there back up to the water spray assembly 14.
The water spray assembly 14 comprises a water box 48 which extends
along the side wall 26 and a pair of distribution pipes 50 which
extend horizontally from the water box across the interior of the
conduit 10 to its opposite wall 28. Each of the pipes 50 is fitted
with a plurality of nozzles 52 which emit mutually intersecting fan
shaped water sprays to provide an even distribution of water over
the entire coil assembly 16.
The mist eliminator assembly 12 comprises a plurality of closely
spaced elongated strips 54 which are bent along their length to
form sinuous paths from the region of the water spray assembly out
through the top of the conduit 10. It will be noted that the mist
eliminator assembly extends across substantially the entire
cross-section of the conduit, and, since the cross-section of the
conduit 10 is substantially uniform, the mist eliminator assembly
occupies substantially the same cross-sectional area of the conduit
10 as the coil assembly 16.
The coil assembly 16 comprises an upper inlet manifold 56 and a
lower outlet manifold 58 which extend horizontally across the
interior of the conduit 10 adjacent the side wall 26. As can be
seen in FIG. 3, the manifolds are held in place by means of
brackets 60 on the side wall 26. Inlet and outlet fluid conduits 62
and 64 extend through the side wall 26 and communicate with the
upper and lower manifolds 56 and 58 respectively. These fluid
conduits are connected to receive a fluid to be cooled or
condensed, for example the refrigerant from a compressor in an air
conditioning system (not shown).
A plurality of cooling tubes 66 are connected between the upper and
lower manifolds 56 and 58. Each tube is formed into a serpentine
arrangement by means of 180.degree. bends 68 and 70 (FIG. 4) near
the side walls 26 and 28 so that different segments of each tube
extend generally horizontally across the interior of the conduit 10
back and forth between the side walls 26 and 28 at different levels
in the conduit along a vertical plane parallel and closely spaced
to the plane of each of the other tubes 66. It will also be noted
that the tubes 66 are arranged in alternately offset arrays with
each tube being located a short distance lower or higher than the
tubes on each side of it. It can be seen in FIG. 4 that each of the
manifolds 56 and 58 is provided with an upper and a lower row of
openings to accept the tubes 66 at these two different levels. In a
preferred embodiment these tubes have an outside diameter of 1.05
inches (2.67 cm). It is also preferred that each 180.degree. bend
have a radius of two and three thirty second inches (5.32 cm) so
that the segments of each tube will be vertically spaced apart from
each other by four and three sixteenths inches (10.64 cm). Further,
the corresponding levels of the segments of adjacent tubes should
be offset vertically from each other by an amount equal to or
greater than the tube diameter; and an offset of two and one tenth
inches (5.33 cm) is preferred.
In order to support the tubes 66 at the bends 68 and 70 there are
provided horizontally extending support rods 72 which are mounted
at the wall 26, between the brackets 60 and, at the wall 28,
between brackets 74.
There are also provided a plurality of vertical spacer rods 76
which extend between the adjacent tubes 66 near the support rods
72. The spacer rods 76 hold the adjacent tubes 66 a short distance
from each other in the lateral direction as can be seen in the
fragmentary perspective view of FIG. 5, and they are held in place
frictionally between the tubes. The spacer rods 76 preferably have
a diameter of 0.240 inches (0.61 cm).
As can be seen in FIG. 6, the coil assembly 16 in cross-section
comprises arrays of tube segments 66a, 66b, 66c and 66d arranged at
different levels or elevations due to the offset arrangement of
adjacent tubes. In addition, the horizontal spacing S between the
tube segments in each level is greater than the diameter of the
tubes. More specifically, as shown, this spacing is equal to the
diameter D of the tube segments plus twice the thickness t of each
of the two spacer rods 76 between the adjacent tubes segments at
each level. This differs from the prior art fully packed coil
arrangement shown in FIG. 7 where no spacer rods are used. As can
be seen in FIG. 7 the horizontal spacing S.sub.1 between adjacent
tube segments at each level is no greater than the tube diameter D.
It can also be seen in FIGS. 5 and 6 that the vertical staggering
of the adjacent sinuously shaped tubes 66 results in a staggering
of the tube segments at adjacent levels, so that the tubes at one
level are essentially centered between tubes at the next higher and
next lower level. It will also be noted that the spacer rods 76
form clearances extending vertically down through the coil assembly
equal in width to their thickness t. The thickness t of each spacer
rod should be an appreciable amount, but not substantially greater
than one half the diameter of the tubes 66. best results have been
obtained when the spacer rod diameter is slightly less than one
fourth of the tube diameter. With this spacer rod arrangement the
tube segments at each level occupy less than fifty percent but not
substantially less than twenty five percent of the coil assembly
cross section and preferably forty percent of the coil assembly
cross section.
In some instances it may be preferred to use tubes of non-circular
cross section. The term "diameter" in such cases is to be
understood as the diametrical distance across the tube cross
section in a horizontal direction.
In operation of the heat exchanger of FIGS. 1-6 a fluid to be
cooled or condensed, such as a refrigerant from an air conditioning
system, flows into the heat exchanger via the inlet conduit 62.
This fluid is then distributed by the upper manifold 56 to the
upper ends of the cooling tubes 66; and its flows down through the
tubes, back and forth across the interior of the conduit 10 at
different levels therein until it reaches the lower manifold 58
where it is collected and transferred out of the heat exchanger via
the outlet conduit 64. As the fluid being cooled flows through the
tubes 66, water is sprayed from the nozzles 52 down over the outer
surfaces of the tubes and air is blown from the fans 32 up between
the tubes. The sprayed water collects in the trough 20 and is
recirculated through the nozzles. The upwardly flowing air passes
through the mist eliminator assembly 12 and exhausts up out of the
system.
During its downward flow through the cooling tubes 66, the fluid
being cooled gives up heat to the walls of the tubes. This heat
passes outwardly through the tube walls to water flowing down over
their outer surface. As the downwardly flowing water encounters the
upwardly moving air, the water gives up heat to the air, both by
sensible heat transfer and by latent heat transfer, i.e. by partial
evaporation. The remaining water falls back down into the trough 20
where it collects for recirculation. As the upwardly moving air
encounters the downwardly flowing water and extracts heat from the
water, the air also entrains a certain amount of water in the form
of droplets which it carries up out from the coil assembly 16 and
up out of the water spray assembly 14. However, as the air passes
through the mist eliminator assembly 12, its flow is changed
rapidly in lateral directions and the liquid droplets carried by
the air become separated from the air and are deposited on the
elements of the mist eliminator. This water then falls back onto
the spray and coil assemblies. Meanwhile the resulting high
humidity, but essentially droplet free, air is exhausted out
through the top of the conduit 10 to the atmosphere.
It is believed that the transfer of heat from the downwardly
flowing water to the upwardly flowing air is enhanced by the high
relative velocity between the water and air because the air shears
the film of water flowing down over each tube. This shearing, it is
believed, promotes heat transfer by increasing water surface area,
by breaking surface tension and by reducing local ambient pressure.
It is also thought that this shearing action becomes effective when
the upward velocity of the air in the vicinity of the tubes is at
least four hundred lineal feet (122 meters) per minute.
As explained in the standard handbook of the American Society of
Heating, Refrigeration and Air Conditioning Engineers, two separate
heat transfer processes are involved in the operation of
evaporative heat exchangers. In the first heat transfer process,
heat from the fluid beintg cooled or condensed passes through the
tube walls to the water flowing over the tubes. In the second
process, heat is transferred from the water flowing over the tubes
to the upwardly flowing air. These two processes are described by
the following equations:
where
q=total heat transferred;
A=total tube surface area;
t.sub.c =fluid temperature in the tubes;
t.sub.s =water temperature outside the tubes;
U.sub.s =heat transfer coefficient--fluid to water;
h.sub.s enthalpy of saturated air at t.sub.s ;
h.sub.l =enthalpy of ambient air; and
U.sub.c =heat transfer coefficient--water to air.
In both heat transfer processes the amount of heat transferred is
directly proportioned to the total tube surface area. Also, in both
processes, the coefficients U.sub.s and U.sub.c are proportioned to
the relative velocities of the fluids. These two criteria indicate
that maximum heat transfer will occur when a large number of
closely spaced tubes are used in the coil assembly since such an
arrangement maximizes tube surface area as well as air flow
velocity in the region of the tubes.
It has been found, however, that heat transfer in a counterflow
evaporative heat exchanger can be improved by arrangements which
are contrary to that indicated by the foregoing heat transfer
equations. That is, the heat transfer in a counterflow evaporative
heat exchanger was found to increase when the number of tubes in
the coil assembly was decreased and when the air flow velocity in
the vicinity of the tubes was also decreased.
The amount by which heat transfer will be affected as the number of
tubes is reduced and as tube spacing is increased can be seen in
the diagram of FIG. 8. In this diagram, heat rejection is plotted
against tube spacing, expressed as a percentage of tube diameter,
for a coil assembly as shown in FIG. 5-7. The different tube
spacings are obtained by removal of tubes from the coil assembly
and repositioning the remaining tubes to maintain the same overall
coil assembly cross section. In the example used the minimum tube
spacing is equal to one tube diameter; and this corresponds to the
spacing S.sub.1 in FIG. 7. Three different curves A, B and C
represent the heat rejection for different flow rates of water
sprayed over the tubes, with curve A corresponding to three gallons
per square foot (122 liters per square meter) of projected area of
coil assembly cross section per minute, curve B corresponding to
four and one half gallons per square foot (183 liters per square
meter) and curve C corresponding to six gallons per square foot
(244 liters per square meter) per minute.
As can be seen in FIG. 8, as the tube spacing is increased from one
hundred percent of tube diameter (prior art), the amount of heat
transfer actually increases up to a maximum where the tube spacing
corresponds to one hundred twenty percent of tube diameter. This
corresponds to a reduction of about twenty percent in the total
tube surface area of the coil assembly; and it also represents a
significant reduction in the cost of the coil assembly. As the tube
spacing is further increased, and the total number of tubes is
correspondingly decreased, the overall heat transfer from the coil
assembly also decreases, but it remains higher than for the closely
packed coil assemblies of the prior art until the tube spacing is
about one hundred thirty percent of the tube diameter. This
corresponds to a reduction of about thirty percent of the total
tube surface area of the tube assembly. Even when the tube spacing
is further increased, the amount of heat transfer per unit area of
cooling tube surface remains higher than for prior art closely
packed coil assemblies. However, the total heat transfer of the
overall coil assembly falls off beyond practical limits when the
number of tube spacing is about two hundred percent of tube
diameter, i.e. when the spacing at each level in the conduit is
about twice the tube diameter. It will be understood that as the
tube spacing is increased, the thickness of the spacer rods 76 is
correspondingly increased.
The upward velocity of the air between the tubes 66 should be at
least four hundred feet (122 meters) per minute, but less than
fourteen hundred feet (427 meters) per minute and, preferably,
about one thousand feet (305 meters) per minute to obtain the
benefits of this invention. It has been found that when air is
blown into the conduit 10 at a velocity of about six hundred feet
(183 meters) per minute, the performance characteristics of FIG. 8
can be expected. It will be appreciated that for a given flow rate
of air into the conduit 10 the velocity of the air in the region of
the tubes will increase in inverse proportion to the amount of
space between the tubes so that in a closely packed coil assembly
the air velocity will be generally higher than in a coil assembly
having spaced apart tubes.
It has been found also that the use of the spaced tube coil
assembly of the present invention makes it possible to obtain
additional improvements in heat transfer by increased rates of
water spray. As can be seen at the extreme right side of the
diagram of FIG. 8, where curves A, B and C essentially merge, the
amount of water sprayed over the closely packed coil assembly of
the prior art does not have a significant effect on heat transfer;
however, where the spaced tube coil assembly of the present
invention is used, heat transfer can be significantly increased by
increasing the amount of water sprayed over the coil assembly. The
use of a large cooling water flow rate provides a still further
advantage in that it improves the washing effect of the cooling
water and reduces scale buildup on the tubes.
While it is not known positively why the spaced tube coil assembly
of this invention provides improved heat transfer, it is believed
that two factors cooperate to bring about this effect.
Firstly, it is thought that the reduced air velocity which results
from the increased tube spacing prevents the air from scrubbing the
downwardly flowing water from the tube surfaces. In this manner the
total tube surface area through which heat can transfer directly to
the downwardly flowing water is maximized. While the upward
velocity of the air between the tubes should be sufficient to
produce a shearing action on the water films flowing over the
tubes, and even an entrainment of droplets which are carried up out
of the coil assembly, the upward velocity of the air should not be
so great that it actually strips the film from the surface of the
tube. It is believed that if the air velocity is too high, the air
will scrub the water film from the tube surface effectively
reducing heat transfer surface area so that heat transfer from the
tube will be impaired. It is also believed that the velocity of the
air in the vicinity of the tubes should be less than fourteen
hundred feet (427 meters) per minute.
The second factor involved in the enhancement of heat transfer in
the system of the present invention is the greater flow velocity
which the fluid being cooled or condensed must undergo in passing
through a reduced number of tubes. In order to accommodate a given
amount of fluid to be cooled with a coil assembly having fewer
cooling tubes than prior art coil assemblies, it is necessary, in
the case of the present invention, for the fluid being cooled to
flow at a higher velocity through the cooling tubes than it did in
prior art closely packed coil assembly tubes. This higher velocity
enhances the heat transfer from the fluid being cooled to the tube
walls.
The foregoing factors are believed to cooperate in combination to
provide a counterflow type heat exchanger with greater heat
transfer capability and lower cost than was obtained by the prior
art.
It is to be understood that regardless of the correctness of the
foregoing explanation, it has been found in actual tests that the
phenomenon of improved heat transfer is obtained when the tube
spacing is maintained such that at each level in the coil assembly
the adjacent tube segments are spaced apart by more than one tube
diameter but not substantially more than two tube diameters, and
when the velocity of the air in the vicinity of the tubes is
maintained at less than fourteen hundred feet (427 meters) per
minute but not substantially less than six hundred feet (183
meters) per minute; and it has found that maximum heat transfer is
obtained when the adjacent tube segments are spaced apart by about
one and one half tube diameters and when the velocity of the air in
the vicinity of the tubes is maintained at about one thousand feet
(305 meters) per minute.
It is to be understood that the present invention does not pertain
to co-current flow heat exchangers wherein the sprayed water and
cooling air both flow in parallel or downwardly past a coil
assembly. In those systems the relative velocity between the air
and the water is not high and the overall heat transfer capability
of such devices is much lower than in counterflow heat exchangers
of similar size. Co-current flow heat exchangers employ coil
assemblies with large spacings between the adjacent tubes for the
same reason that prior art countercurrent flow heat exchangers
employ coil assemblies with small spacings between the adjacent
tubes, namely, to increase the relative velocity between the air
and the water by allowing the air to move more freely over the
water without carrying the water along with it. In this invention
however, the tube spacing in a countercurrent heat exchanger is
increased in order to reduce the velocity of the air moving up
against the downwardly flowing water, which is precisely opposite
to the purpose of spacing tubes in prior art co-current flow
evaporation heat exchangers.
The present invention is also not concerned with heat exchangers,
even of the counterflow type, in which air velocities are so low
that the upwardly flowing air did not entrain any appreciable
amount of water. In those devices no substantial amount of heat
transfer was obtained and if any mist eliminator was needed at all,
it would only be employed where the air exhaust was through a very
small opening which produced high air exit velocities far greater
than the air velocity over the cooling tubes. In the case of the
present invention, air velocities in the region of one thousand
feet (305 meters) per minute are employed in the region of the
cooling tubes and accordingly in order to enable the entrained
water to be removed from the air the mist eliminator assembly 12
should extend over substantially the same cross-sectional area as
the coil assembly 16. In this manner the air velocity in the region
of the mist eliminator assembly will not be appreciably higher than
in the region of the coil assembly and the mist eliminator assembly
will be effective to remove the majority of the entrained water
from the exiting air.
FIG. 9 shows a modified version of the present invention. The heat
exchanger shown in FIG. 9 is the same as that of FIGS. 1-6 in all
respects except that in FIG. 9 there is provided a propeller
assembly 118 which replaces the fan assembly 18 of the preceding
embodiment. The propeller assembly 118 blows air into the conduit
10 via a cowl 134 in a manner similar to the centrifugal fans 32.
The propeller assembly 118 is capable of moving as large a quantity
of air as the centrifugal fan 32 but with substantially less power
than is required by the centrifugal fan. In order for a propeller
to operate efficiently to move large quantities of air, however, it
is important that the static pressure difference between the
propeller input and output be minimized. With the open or spaced
tube coil assembly of the present invention the pressure drop
across the coil is minimized and accordingly it becomes possible
with the present invention to employ a propeller drive for the
cooling air in a very efficient manner.
It has also been found that the spaced tube coil assembly of the
present invention, with its vertical clearances between adjacent
tubes provides access for tools and cleaning implements to all tube
surfaces and thereby maintenance of the coil assembly is
facilitated.
Having thus described the invention with particular reference to
the preferred forms thereof, it will be obvious to those skilled in
the art to which the invention pertains, after understanding the
invention, that various changes and modifications may be made
therein without departing from the spirit and scope of the
invention as defined by the claims appended hereto.
* * * * *